Vehicle drive system

The disclosure of Japanese Patent Application No. 2008-145669 filed on Jun. 3, 2008, including the specification, drawings and abstract is incorporated herein by reference in its entirety.

Background of the invention

1. Field of the invention

The invention relates to a vehicle drive system in which an internal combustion engine and an electric motor are provided as driving sources.

2. Description of the related art

In a conventional vehicle drive system (as disclosed in Japanese Patent Application Publication No. 2002-204504 (JP-A-2002-204504)), a twin-clutch type transmission using a plurality of gears is coupled to an internal combustion engine, and an electric motor is disposed between two clutch shafts of the transmission, such that a rotor shaft of the motor extends in a direction perpendicular to the clutch shafts. One of the clutch shafts is connected to the rotor shaft of the motor via a pair of bevel gears, and the other clutch shaft is connected to a case of the motor via another pair of bevel gears. In operation, a difference between the rotational speeds of the two clutch shafts is utilized to drive the motor, thereby to generate electric power. Other prior art documents relevant to the invention include, for example, Japanese Patent Application Publication No. 2004-293795 (JP-A-2004-293795).

In the drive system of JP-A-2002-204504, the internal combustion engine and the two clutch shafts are connected to each other via two or more coupling gears interposed therebetween. Therefore, the drive system inevitably suffers from losses due to engagement of the gears during transmission of power of the engine. Also, it is difficult to reduce the size of the drive system since the electric motor is disposed between the two clutch shafts. Furthermore, the rotor shaft of the motor and each of the clutch shafts, which are perpendicular to each other, are connected to each other via bevel gears, and this arrangement makes the drive system complicated in construction.

Summary of the invention

The present invention provides a drive system operable to place a transmission mechanism in a direct-drive gear position by connecting an internal combustion engine with an output member without involving any change gear train therebetween, thereby to reduce losses due to engagement of gears.

According to a first aspect of the invention, there is provided a drive system of a vehicle, which includes: an internal combustion engine, an electric motor, an output member that delivers power to driving wheels of the vehicle, a differential mechanism that includes a first element, a second element and a third element capable of differential rotation relative to each other, the internal combustion engine being coupled to the first element, the electric motor being connected to the second element, a transmission mechanism that establishes a plurality of gear positions that are set stepwise in decreasing order of a gear ratio over a range from the differential mechanism to the output member, by switching among change gear trains interposed between the first element or the third element of the differential mechanism and the output member for transmitting power from the first element or the third element to the output member, and coupling the first element to the output member without involving any of the change gear trains, and a differential lock mechanism that switches the differential mechanism between a locked state in which selected two of the first element, the second element and the third element of the differential mechanism are coupled to each other so as to be inhibited from differential rotation, and a released state in which the inhibition of the differential rotation is cancelled.

According to the drive system as described above, the differential lock mechanism switches the differential mechanism between the locked state and the released state while the operation of the electric motor is being controlled. With this arrangement, during shifting (change of gear positions) of the transmission mechanism through switching from a change gear train interposed between the first element and the output member to a change gear train interposed between the third element and the output member, for example, these change gear trains can be easily brought into synchronization. Therefore, the gear position can be changed with reduced shift shock. In addition to the gear positions established using change gear trains, the transmission mechanism is operable to establish a direct-drive gear position in which the first element to which the engine is coupled is coupled directly to the output member, with no change gear train interposed therebetween. This arrangement leads to reduction of losses due to engagement of gears, and improvement of the fuel economy or efficiency.

In the drive system as described above, the gear positions of the transmission mechanism may include a direct-drive gear position that is established by coupling the first element to the output member without involving any of the change gear trains, and an overdrive gear position that has a smaller gear ratio than that of the direct-drive gear position, and a shift control device may be further provided for controlling the transmission mechanism so as to shift the transmission mechanism from the overdrive gear position to the direct-drive gear position when the vehicle speed becomes equal to or higher than a specified value while the overdrive gear position is established.

For example, when the vehicle speed is in a high-vehicle-speed range equal to or higher than 160 km/h, the engine optimum operating line is located in a high-engine-speed region, and therefore the optimum gear ratio is slightly larger (low-gear side) than the gear ratio of the overdrive gear position suitable for vehicle speeds around 120 km/h. According to the drive system as described above, when the vehicle speed becomes equal to or higher than the specified value while the overdrive gear position is established, the transmission mechanism is shifted into the direct-drive gear position having the larger gear ratio than the overdrive gear position. Namely, the transmission mechanism is downshifted from the overdrive gear position to the direct-drive gear position. As a result, the vehicle is driven at a gear ratio or speed ratio suitable for the high-vehicle-speed range equal to or higher than the specified value, thus assuring improved fuel economy. Also, the gear position into which the transmission mechanism is shifted in the high-vehicle-speed range is the direct-drive gear position involving no change gear train; therefore, losses due to engagement of gears can be reduced. Since the losses due to engagement of gears increase in proportion to the torque, the greatest advantage can be provided by selecting the direct-drive gear position in the high-vehicle-speed range.

In the drive system as described above, the transmission mechanism may include a first driveshaft that rotates with the first element, a hollow second driveshaft that is disposed coaxially around the first driveshaft and rotates with the third element, and a countershaft that is disposed in parallel with the first driveshaft and the second driveshaft, and is provided with a counter gear for transmitting power to the output member, and one of the change gear trains that provides one of the gear positions having the largest gear ratio may be provided between the first driveshaft and the countershaft.

Where the distance between the driveshaft and the countershaft is fixed, the gear ratio of a certain change gear train can be increased with reduction in the gear radius of a drive gear that constitutes the change gear train. In the drive system as described above, the drive gear of the change gear train having the largest gear ratio is mounted on the first driveshaft as the inner one of the coaxially assembled driveshafts; therefore, the gear radius can be reduced as compared with the case where the drive gear is mounted on the second driveshaft as the outer shaft. Thus, the distance between the driveshaft and the countershaft can be reduced as compared with the case where the change gear train is provided between the outer second driveshaft and the countershaft so as to provide a desired gear ratio. Consequently, the size of the drive system, in particular, the radial dimension of the drive system, can be reduced.

In the drive system as described above, the transmission mechanism may include a first driveshaft that rotates with the first element, a hollow second driveshaft that is disposed coaxially around the first driveshaft and rotates with the third element, countershafts that are disposed in parallel with the first driveshaft and is provided with counter gears for transmitting power to the output member, and clutches that are disposed radially outwardly of the second driveshaft and selectively couples the change gear trains that provide the above-indicated plurality of gear positions, to the countershafts, and the differential lock mechanism may place the differential mechanism in the locked state by coupling the first driveshaft and the second driveshaft to each other, and may be disposed between the first driveshaft and one of the clutches.

According to the drive system as described above, the differential lock mechanism and the clutch are positioned so as to overlap each other in the axial direction. Therefore, the dimension of the drive system as measured in the axial direction is prevented from increasing due to the addition of the differential lock mechanism.

In the drive system as described above, the electric motor may include a stator, and a rotor that is disposed coaxially with the stator and has a space formed radially inside thereof, and the differential mechanism may be placed in the space formed radially inside the rotor, while the differential lock mechanism may be located at a position remote from the differential mechanism in an axial direction.

In this case, since the differential mechanism is placed in the space formed radially inside the rotor, the differential mechanism and the motor are positioned so as to overlap each other in the axial direction. If the differential lock mechanism were located in the vicinity of the differential mechanism, the arrangement around the motor would be complicated. In this case, it would be difficult to position the motor and the differential mechanism to overlap each other, and the motor and/or the differential mechanism would need to be displaced in the axial direction. Where the drive system is employed in a FR vehicle in which the axial direction of the system coincides with the running direction of the vehicle, it is advantageous to install the electric motor as closest to the front of the vehicle as possible, in view of the shape of the floor tunnel whose interior space is narrowed toward the rear of the vehicle. This is because the torque produced by the motor increases in proportion to the second to third power of the radius of the motor. If the radius of the motor is reduced, the dimension of the motor as measured in the axial direction must be increased so as to provide the same torque as that provided when the motor radius is not reduced. When the drive system as described above, in which the differential lock mechanism is located remote from the differential mechanism, is employed in the FR vehicle, the electric motor can be placed in a front part of the vehicle. Accordingly, the radius of the motor may be increased under reduced restrictions, and the axial dimension of the motor may be reduced.

In the drive system as described above, the transmission mechanism may include a driveshaft to which power is transmitted from the differential mechanism, and first countershaft and second countershaft that are disposed in parallel with the driveshaft and transmit power of the driveshaft to the output member, and a first change gear train provided between the driveshaft and the second countershaft for establishing a gear position having the smallest gear ratio, and a second change gear train that is provided between the driveshaft and the first countershaft for establishing a gear position having a gear ratio that is larger by two steps than that of the first change gear train may be aligned with each other in a direction perpendicular to the axial direction of the driveshaft.

In the drive system as described above, the first change gear train and the second change gear train are positioned so as to overlap each other in the axial direction. Therefore, the dimension of the drive system as measured in the axial direction can be reduced. Also, the first gear train that provides the gear position having the smallest gear ratio and the second gear train that provides the gear position having a gear ratio that is larger by two steps than the smallest gear ratio are aligned with each other in radial directions (i.e. in a direction perpendicular to the axial direction of the driveshaft); therefore, a difference in the gear ratio between the aligned gear trains may be smaller than that in the case where gear trains that provide other gear positions are aligned with each other in radial directions. Consequently, the distance between the first countershaft and the second countershaft can be reduced.

In the drive system as described above, the transmission mechanism may include a driveshaft to which power is transmitted from the differential mechanism, and a countershaft that is disposed in parallel with the driveshaft and transmit power of the driveshaft to the output member, and a change gear train as one of the change gear trains which is disposed between the driveshaft and the countershaft and provides a gear position having the largest gear ratio may be located at a position remotest from the internal combustion engine.

The torque applied to the change gear train that provides a gear position having the largest gear ratio is larger than those applied to the other gear trains. In the drive system as described above, the change gear train that provides the gear position having the largest gear ratio is located farther apart from the engine than the other gear trains, and the change gear train having the largest gear ratio is supported at a position close to one end of the driveshaft. Accordingly, deflection of the driveshaft can be suppressed, as compared with the case where the change gear train having the largest gear ratio is supported at a position close to the middle of the driveshaft.

As explained above, according to the present invention, the transmission mechanism establishes the direct-drive gear position by coupling the first element to which the engine is coupled, directly to the output member, without involving any change gear train, in addition to the gear positions involving the respective change gear trains; therefore, losses due to engagement of gears can be reduced.

Brief description of the drawings

The features, advantages, and technical and industrial significance of this invention will be described in the following detailed description of example embodiments of the invention with reference to the accompanying drawings, in which like numerals denote like elements, and wherein:

FIG. 1 is a view schematically showing a vehicle in which a drive system according to a first embodiment of the invention is employed;

FIG. 2 is a view showing a clutch engagement table indicating correspondences between the operating patterns of the drive system and the operating state of each clutch;

FIG. 3 is a view showing a power transmission path established in a first-gear EV mode in the first embodiment;

FIG. 4 is a view showing one example of alignment chart showing the operation of the drive system of the first embodiment in the first-gear EV mode;

FIG. 5 is a view showing a power transmission path established in a first-gear Eng mode in the first embodiment;

FIG. 6 is a view showing one example of alignment chart showing the operation of the drive system of the first embodiment in the first-gear Eng mode;

FIG. 7 is a view showing a power transmission path established in a first-gear Eng+MG mode in the first embodiment;

FIG. 8 is a view showing one example of alignment chart showing the operation of the drive system of the first embodiment in the first-gear Eng+MG mode;

FIG. 9 is a view showing a power transmission path established in a second-gear CVT mode in the first embodiment;

FIG. 10 is a view showing one example of alignment chart showing the operation of the drive system of the first embodiment in the second-gear CVT mode;

FIG. 11A is an alignment chart showing a first condition during shifting from first gear to second gear in the Eng+MG mode;

FIG. 11B is an alignment chart showing a second condition during shifting from first gear to second gear in the Eng+MG mode;

FIG. 11C is an alignment chart showing a third condition during shifting from first gear to second gear in the Eng+MG mode;

FIG. 11D is an alignment chart showing a fourth condition during shifting from first gear to second gear in the Eng+MG mode;

FIG. 11E is an alignment chart showing a fifth condition during shifting from first gear to second gear in the Eng+MG mode;

FIG. 12 is an explanatory view showing an engine optimum operating line, maximum torque and an equal-power line, in which the vertical axis indicates engine torque, and the horizontal axis indicates the engine speed;

FIG. 13 is a view showing one example of alignment chart showing the operation of the drive system of the first embodiment in the sixth-gear CVT mode;

FIG. 14 is an explanatory view showing an engine optimum operating line, engine operating line (point) in sixth gear, and an engine operating line (point) in fifth gear, in which the vertical axis indicates engine torque, and the horizontal axis indicates the engine speed;

FIG. 15 is a view schematically showing a vehicle in which a drive system according to a second embodiment of the invention is employed;

FIG. 16 is a view showing a clutch engagement table of the drive system shown in FIG. 15;

FIG. 17 is a view schematically showing a vehicle in which a drive system according to a third embodiment of the invention is employed; and

FIG. 18 is a view showing a clutch engagement table of the drive system shown in FIG. 17.

Detailed description of embodiments

FIG. 1 schematically illustrates a vehicle in which a drive system according to a first embodiment of the invention is employed. As shown in FIG. 1, the vehicle 1 is a so-called hybrid vehicle having a FR (front-engine rear-drive) type layout. The vehicle 1 is provided with a drive system 2A for running the vehicle. The drive system 2A includes an internal combustion engine 3, a motor-generator (MG) 4 as an electric motor, an output shaft 5 as an output member that delivers power to driving wheels 9 of the vehicle 1, a power distribution mechanism 6 as a differential mechanism to which the engine 3 and the MG4 are coupled, and a transmission mechanism 7 that changes the speed of the output of the power distribution mechanism 6 and transmits the resulting power output to the output shaft 5. The rotation of the output shaft 5 is transmitted to the left and right driving wheels 9, respectively, via a propeller shaft 12 and a differential gear 13. The propeller shaft 12 is coupled to the output shaft 5 and the differential gear 13, respectively, via universal joints (not shown) provided at opposite ends of the propeller shaft 12.

The internal combustion engine 3 is in the form of a spark ignition type, multi-cylinder engine. The engine 3 is connected to a first driveshaft 10A via an engine clutch SE that allows or inhibits power transmission therebetween. The engine clutch SE is in the form of a conventional dog clutch, and is operable to switch between an engaged state in which power is transmitted from the engine 3 to the first driveshaft 10A, and a released state in which the power transmission is cut off. The first driveshaft 10A extends through the MG 4 and the power distribution mechanism 6, and is disposed on the same axis as the output shaft 5. A hollow second driveshaft 10B is disposed around the first driveshaft 10A, coaxially with the first driveshaft 10A. The engine 3 is provided with a starter generator (SG) 11 operable to start the engine 3, and generate electric power using power of the engine 3 after it is started. The electric power generated by the SG 11 is consumed at various parts of the vehicle 1, and is also stored in a battery (not shown) provided as a power supply of the MG 4.

The MG 4 is arranged to function as an electric motor and also function as a generator. The battery is electrically connected to the MG4 via an inverter (not shown), and the inverter is controlled so as to set output torque or regenerative torque of the MG 4 appropriately. The MG 4 includes a rotor 4a that is rotatably provided coaxially with the first driveshaft 10A, and a stator 4b located radially outwardly of the rotor 4a. The stator 4b is fixed to a stationary element, such as a casing (not shown), so as not to rotate. The power distribution mechanism 6 is placed in a space formed radially inside the rotor 4a. Namely, the MG 4 and the power distribution mechanism 6 are positioned so as to overlap each other as viewed in the direction of the axis of the first driveshaft 10A.

The power distribution mechanism 6 is in the form of a planetary gear set. More specifically, the power distribution mechanism 6 includes a sun gear Sn as an external gear, a ring gear Rg as an internal gear disposed coaxially with the sun gear Sn, and a carrier Cr that holds pinions 15 engaging with the sun gear Sn and ring gear Rg such that the pinions 15 can rotate about themselves and about the axis of the first driveshaft 10A. Thus, the power distribution mechanism 6 is a conventional gear mechanism in which the above three elements Sn, Rg and Cr can rotate at different speeds in different directions from each other. In this embodiment, the engine 3 is connected to the carrier Cr via the first driveshaft 10A, and the MG4 is coupled to the sun gear Sn, while the second driveshaft 10B is coupled to the ring gear Rg. With this arrangement, the first driveshaft 10A rotates integrally with the carrier Cr, and the second driveshaft 10B rotates integrally with the ring gear Rg. Thus, the carrier Cr is one example of first element according to the invention, and the sun gear Sn is one example of second element according to the invention, while the ring gear Rg is one example of third element according to the invention.

The transmission mechanism 7 is placed in one of a plurality of gear positions, by selecting one of change gear trains interposed between the driveshaft 10A or 10B and the output shaft 5, or coupling the first driveshaft 10A to the output shaft 5 without involving any change gear train. The transmission mechanism 7 provides a plurality of gear positions comprising six forward-drive gear positions and one rear-drive gear position. The six forward-drive gear positions are first-speed, second-speed, third-speed, forth-speed, fifth-speed and sixth-speed gear positions (which may also be called "first gear", "second gear", "third gear", "fourth gear", "fifth gear" and "sixth gear"), which are set stepwise in decreasing order of the gear ratio or speed ratio over a range from the power distribution mechanism 6 to the output shaft 5. Of these gear positions, the fifth-speed gear position is a direct-drive position established by coupling the first driveshaft 10A directly to the output shaft 5 without involving any change gear train, and the gear ratio of this position is equal to 1.

The first driveshaft 10A protrudes from an end portion of the second driveshaft 10B, and a 1st-speed/reverse drive gear 16 for shared use in first gear and reverse gear and a 3rd-speed drive gear 17 for use in third gear are mounted on a portion of the first driveshaft 10A which protrudes from the second driveshaft 10B. Also, a 2nd-speed drive gear 18 for use in second gear and a 4th-speed/6th-speed drive gear 19 for shared use in fourth gear and sixth gear are mounted on the second driveshaft 10B.

The transmission mechanism 7 includes two countershafts 20A, 20B disposed in parallel with the driveshafts 10A, 10B. Driven gears 21-24 that are in mesh with the respective drive gears 16-19 are rotatably mounted on the first countershaft 20A. A first counter gear 31 that is in mesh with an output gear 30 that rotates with the output shaft 5 is mounted on an end portion of the first countershaft 20A, such that the counter gear 31 can rotate integrally with the countershaft 20A. Also, a driven gear 26 that is in mesh with the drive gear 19 mounted on the second driveshaft 10B, and a driven gear 29 for use in reverse gear are mounted on the second countershaft 20B. The reverse driven gear 29 engages with the drive gear 16 mounted on the first driveshaft 10A, via an intermediate gear 33 interposed therebetween. A second counter gear 32 that is in mesh with the output gear 30 is mounted on an end portion of the second countershaft 20B, such that the counter gear 32 can rotate integrally with the countershaft 20B.

Suitable combinations of the gears mounted on the driveshafts 10A, 10B and the countershafts 20A, 20B engage with each other so as to provide a plurality of change gear trains having different gear ratios. More specifically, a change gear train 41 that provides the first-speed gear position consists of the gears 16, 21 provided between the first driveshaft 10 and the first countershaft 20A, and a change gear train 42 that provides the second-speed gear position consists of the gears 18, 22 provided between the second driveshaft 10B and the first countershaft 20A, while a change gear train 43 that provides the third-speed gear position consists of the gears 17, 23 provided between the first driveshaft 10A and the first countershaft 20A, and a change gear train 44 that provides the fourth-speed gear position consists of the gears 19, 24 provided between the second driveshaft 10B and the first countershaft 20A. A change gear train 46 that provides the sixth-speed gear position consists of the gears 19, 26 provided between the second driveshaft 10B and the second countershaft 20B, and a change gear train 50 that provides reverse-drive gear position consists of the gears 16, 33, 29 provided between the first driveshaft 10A and the second countershaft 20B.

While the first driveshaft 10A and the second driveshaft 10B are assembled coaxially with each other, the 1st-speed drive gear 16 having the largest gear ratio is mounted on the first driveshaft 10A as the inner shaft. Since the first-speed gear position needs to provide the largest gear ratio, the 1st-speed drive gear 16 needs to be formed with the smallest gear radius. If the 1st-speed drive gear 16 were mounted on the second driveshaft 20B as the outer shaft, the gear radius of the drive gear 16 would be naturally large. In this case, the gear radius of the driven gear 21 that meshes with the drive gear 16 would also be increased so as to provide a desired gear ratio, resulting in an increase in the distance between the first driveshaft 10A and the first countershaft 20A. In this embodiment in which the 1st-speed drive gear 16 is mounted on the inner, first driveshaft 10A, the distance between the axis of the first driveshaft 10A and that of the first countershaft 20A can be reduced as compared with the case where the drive gear 16 is mounted on the second driveshaft 20B, thus achieving reduction of the radial dimension or size of the drive system 2A. Consequently, the possibility of successful installation of the drive system 2A on the vehicle is increased, and the weight of the drive system 2A may be reduced. In the above-described case, the first countershaft 20A is one example of countershaft according to the invention.

The drive gear 19 mounted on the second driveshaft 10B is shared to establish the fourth-speed gear position and the six-speed gear position, and the 4th-speed change gear train 44 and the 6th-speed change gear train 46 are aligned with each other in radial directions. Namely, the change gear train 44 and the change gear train 46 are positioned so as to overlap each other as viewed in the axial direction. Since the drive gear 19 is used in both of the fourth-speed and sixth-speed gear positions, the dimension of the drive system 2A as measured in the axial direction can be reduced. Also, since the drive gear 19 is shared between the sixth-speed gear position having the smallest gear ratio and the fourth-speed gear position having a gear ratio that is larger by two steps than that of the sixth-speed gear position, the distance between the axis of the first countershaft 20A and that of the second countershaft 20B can be reduced as compared with the case where the driver gear is shared between gear positions whose gear ratios are different by three steps or more from each other. In this case, the 6th-speed change gear train 46 is one example of first change gear train according to the invention, and the 4th-speed change gear train 44 is one example of second change gear train according to the invention. Also, the second driveshaft 10B is one example of driveshaft according to the invention.

The change gear train 50 that provides the reverse-drive gear position is positioned so as to overlap the change gear train 41 that provides the first-speed gear position in the axial direction, and the drive gear 16 is included in both the change gear train 50 and the change gear train 41. The drive gear 16 for use in first gear has the smallest gear radius. Therefore, the gear radius of the driven gear 29 for use in reverse gear can be made relatively large. As a result, the speed reduction ratio of the reverse-drive gear position can be made large within a limited space.

The second countershaft 20B is shared between the change gear train 46 for the sixth-speed gear position and the change gear train 50 for the reverse-drive gear position, and is formed as another shaft, independently of or separately from the first countershaft 20A on which the change gear trains 41-44 for the other gear positions are mounted. With this arrangement, the gear ratios can be set individually for the respective countershafts. For the reverse-drive gear position and the sixth-speed gear position, for example, the necessity to take account of continuity across the gear ratios in view of the driveability is low. Accordingly, the gear ratios of these gear positions can be set independently of or irrespective of the gear ratios of the other gear positions; therefore, the gear ratio of the sixth-speed gear position having the smallest gear ratio, for example, can be set more freely, i.e., within an increased range. Thus, the setting of the optimum gear ratio, for improvement of the fuel economy, can be easily achieved. For the gear positions of which continuity across the gear ratios are deemed important, the change gear trains that provide these gear positions are mounted on the same first countershaft 20A, and therefore, the gear ratios of the respective gear positions can be set freely in view of the driveability.

The transmission mechanism 7 includes first clutch S1, second clutch S2 and fourth clutch S4 for selectively coupling the driven gears 21-24, 26, 29 to a corresponding one of the countershafts 20A, 20B, so that power is transmitted from the power distribution mechanism 6 to the output shaft 5, via a selected one of the change gear trains 41-46, 50 interposed between the power distribution mechanism 6 and the output shaft 5. The transmission mechanism 7 also includes a third clutch S3 adapted to connect the first driveshaft 10A with the output shaft 5, so as to establish the fifth-speed gear position as the direct-drive position, or disconnect the first driveshaft 10A from the outer shaft 5. Each of the clutches S1-S4 takes the form of a conventional dog clutch. The first clutch S1 is disposed between the 1st-speed driven gear 21 and the 3rd-speed driven gear 23, and the second clutch S2 is disposed between the 2nd-speed driven gear 22 and the 4th-speed driven gear 24, while the third clutch S3 is disposed between the first driveshaft 10A and the output shaft 5, and the fourth clutch S4 is disposed between the 6th-speed driven gear 26 and the reverse-drive driven gear 29. Since the fourth clutch S4 is shared between the sixth-speed gear position and the reverse-drive gear position, the axial dimension of the drive system 2A is made smaller than that in the case where a clutch exclusively used for reverse gear is additionally provided.

The first clutch S1 includes a hub 51 formed integrally with the first countershaft 51, a sleeve 52 that is fitted via splines on the hub 51 to be movable in the axial direction, splines 53 formed integrally with the 1st-speed driven gear 21 and engageable with splines of the sleeve 52, and splines 54 formed integrally with the 3rd-speed driven gear 23 and engageable with the splines of the sleeve 52. With this arrangement, the sleeve 52 may be moved to the 1st-speed driven gear 21 for engagement with the splines 53, so that the driven gear 21 is connected to the first countershaft 20A to be rotatable with the countershaft 20A. As a result, the power of the first driveshaft 10A is transmitted to the output shaft 5 via the change gear train 41, so as to establish the first-speed gear position having the largest gear ratio. Also, the sleeve 52 of the first clutch S1 may be moved to the 3rd-speed driven gear 23 for engagement with the splines 54, so that the driven gear 23 is connected to the first countershaft 20A to be rotatable with the countershaft 20A. As a result, the power of the first driveshaft 10A is transmitted to the output shaft 5 via the change gear train 43, so as to establish the third-speed gear position. Thus, the first clutch S1 is selectively operable among an operating position P1 for establishing the first-speed gear position, an operating position P3 for establishing the third-speed gear position, and a neutral position N in which the driven gears 21, 23 are disconnected from the first countershaft 20A. Thus, the first clutch S1 can selectively establish the first-speed gear position and the third-speed gear position.

Similarly, the second clutch S2 for selectively establishing the second-speed gear position and the fourth-speed gear position includes a hub 55 formed integrally with the first countershaft 20A, a sleeve 56 that is fitted via splines on the hub 55, splines 57 formed integrally with the 2nd-speed driven gear 22, and splines 58 formed integrally with the 4th-speed driven gear 24. With this arrangement, the second clutch S2 is selectively operable among an operating position P2 for establishing the second-speed gear position, an operating position P4 for establishing the fourth-speed gear position, and a neutral position N in which the driven gears 22, 24 are disconnected from the first countershaft 20A. The fourth clutch S4 for selectively establishing the sixth-speed gear position and the reverse-drive gear position includes a hub 60 formed integrally with the second countershaft 20B, a sleeve 61 that is fitted via splines on the hub 60, splines 62 formed integrally with the 6th-speed driven gear 26, and splines 63 formed integrally with the reverse-drive driven gear 29. With this arrangement, the fourth clutch S4 is selectively operable among an operating position P6 for establishing the sixth-speed gear position, an operating position PR for establishing the reverse-drive gear position, and a neutral position N in which the driven gears 26, 29 are disconnected from the second countershaft 20B.

The third clutch S3 for establishing the fifth-speed gear position as the direct-drive position includes a hub 65 formed integrally with the output shaft 5, a sleeve 66 that is fitted via splines on the hub 65 to be movable in the axial direction relative to the hub 65, and splines 67 formed integrally with the first driveshaft 10A and engageable with splines of the sleeve 66. With this arrangement, the sleeve 66 may be moved toward the first driveshaft 10A to be engaged with the splines 67, so that the first driveshaft 10A and the output shaft 5 are coupled to each other to be rotatable as a unit. The sleeve 66 may also be moved to the position illustrated in FIG. 1, so that the first driveshaft 10A and the output shaft 5 are disconnected from each other. Namely, the third clutch S3 is operable between an operating position P5 for establishing the fifth-speed gear position, and a neutral position N in which the first driveshaft 10A and the output shaft 5 are disconnected from each other.

The drive system 2A is provided with a lock clutch SL as a differential lock mechanism for inhibiting differential rotation of the elements of the power distribution mechanism 6. The lock clutch SL is operable to switch the power distribution mechanism 6 between a locked state in which the first driveshaft 10A and the second driveshaft 10B are coupled to each other to be rotatable as a unit so that differential rotation of the power distribution mechanism 6 is inhibited, and a released state in which differential rotation of the power distribution mechanism 6 is allowed. When the first driveshaft 10A and the second driveshaft 10B are coupled to each other by means of the lock clutch SL, the carrier Cr and ring gear Rg of the power distribution mechanism 6, which are coupled to the first driveshaft 10A and the second driveshaft 10B, respectively, rotate at the same speed, and therefore, differential rotation of the elements, i.e., sun gear Sn, ring gear Rg and carrier Cr, is inhibited.

The lock clutch SL takes the form of a conventional dog clutch, and includes a hub 70 formed integrally with the first driveshaft 10A, a sleeve 71 that is fitted via splines on the hub 70 to be movable in the axial direction relative to the hub 70, and splines 72 formed integrally with the second driveshaft 10B and engageable with splines of the sleeve 71. With this arrangement, the sleeve 71 of the lock clutch SL may be moved toward the second driveshaft 10B to be engaged with the splines 72, so that the first driveshaft 10A and the second driveshaft 10B are coupled to each other to be rotatable as a unit, and the power distribution mechanism 6 is switched to the locked state. The sleeve 71 may also be moved to the position illustrated in FIG. 1, so that the first driveshaft 10A and the second driveshaft 10B are disconnected from each other, and the power distribution mechanism 6 is switched to the released state. Namely, the lock clutch SL is operable between an operating position PL for placing the power distribution mechanism 6 in the locked state, and a neutral position N for placing the power distribution mechanism 6 in the released state.

The description continues in the full USPTO document.